Variable nozzle turbine



3 Sheets-sheaf. 1

FIG. I

62 FIG. 2

IN VENTOR.

ATTORNEYS JUDSON s. SWEAR/NGEN Feb. 17, 1970 I J s. SWEARINGEN vARf1ABLEifiozzLE TURBINE Filed Dec. 11. 1967 V r o i. vii" Feb. 17, 1970 J- 9SWEARINGEN 3,495,921

VARIABLE NOZZLE TURBINE Filed Dec. 11, 19s? 3 Sheets- -Shet 2 JUDSON S.SWEARINGEN INVENTOR.

f ,W* g ATTORNEYS Feb. 17, 1970 J. s. SWEARINGEN 3,

. VARIABLE NOZZLE TURBINE Filed Dec. 11. 1967 s Sheets-$heet s V 80 Q7 4v mnnnnrl v JUDSON s. SWEAR/NGEN I N VEN TOR.

BY M ATTORNEYS United States Patent M 3,495,921 VARIABLE NOZZLE TURBINEJudson S. Swearingen, 500 Bel Air Road, Los Angeles, Calif. 90024 FiledDec. 11, 1967, Ser. No. 689,397 Int. Cl. F01d 17/08, 17/00 US. Cl.415163 18 Claims ABSTRACT OF THE DISCLOSURE The turbine disclosed has avariable nozzle assembly which is formed of a plurality of nozzle bladessandwiched between a pair of axially-spaced concentricallymounted rings.The blades are pivotally attached to the two rings. One of the rings isfixed and the other ring, which is an actuator ring, is sealinglymounted on a bearing ring for rotation relative to the fixed ring.Rotation of the actuator ring alters the spacing between complementaryportions of adjacent blades. The blades are so spaced that suchcomplementary portions form the passages from the turbine inlet to theturbine wheel.

There is a pressure drop through the nozzles and the outer face of theactuator ring upstream of the seal is subject to high pressure, anddownstream of the seal is subject to low pressure. The upstream portionof the blade face of the actuator ring is also subject to high pressureand the downstream side of the blade face of the cam ring is subject tolow pressure. However, ordinarily, changes in position of the bladeswill change the location of respective demarcation line dividing thehigh and low pressure areas on the blade side of the actuator ring, suchline being hereinafter called the pressure-break line. Means areprovided in this disclosure to control the pres sure break-line so thatbetween the closed and open position of the blades it will remain inapproximately the same position. Such means are illustrated as being atapered surface on the blade side of the actuator ring so that theactuator ring and the blades are in sealing contact only along a circleadjacent and upstream of the pivot pins mounting the blades. Therefore,a major portion of the upstream side of the actuator ring which isshielded by the blades is relieved so that the area exposed to the highpressure will remain approximately the same regardless of the positionof the blades. To assure balanced forces on both sides of the nozzleblades, the fixed ring may also be tapered in a similar fashion. Toaccomplish full balance, the rings are tapered downstream of thepressure-break line as well as upstream.

Alternate constructions disclosed a nozzle blade which forms an acuteangle with the axis of the turbine wheel and a split actuator ringhaving provisions to prevent over-extension.

BACKGROUND OF THE INVENTION This invention relates to turbines and, moreparticularly, to turbines having variable nozzle assemblies.

In one type of a variable nozzle turbine there is a plurality oftiltable blades sandwiched between a pair of axially-spaced,concentrically-mounted rings. Complementary portions of adjacent bladesform the nozzle openings. Each blade is pivoted on a pin located in astationary ring. Each blade also has an elongated slot which is engagedwith a cam pin located in an opposing, rotatable actuator ring. Rotationof the actuator ring causes the pins in the elongated slots to vary theposition of the blades, thereby changing the size of the nozzleopenings.

If the actuator ring is axially fixed, there is either an excessiveclearance or else a danger of jamming. For example, a clearance of .001inch will permit about a 3,495,921 Patented Feb. 17, 1970 1% blow-bywhile at the same time such a clearance is not sufficient to preventjamming. Therefore, it is preferable that the actuator ring be free tomove axially. With such construction, there is no necessity forclearance because, while the actuator ring is urged against the bladesby pressure to effect a seal, the actuator ring can move away to preventjamming.

There is a pressure drop through the nozzles and high pressure actsagainst the outer face of the actuator ring upstream of the sealdiameter and low pressure acts against the face downstream of the sealdiameter. Both, together, constitute a force which urges the actuatorring against the nozzle blades. An opposing force prevails on the nozzleblade side of the actuator ring. The force on the blade side of theactuator ring is usually somewhat less than the force against the outerface so that the actuator ring is urged gently against the blades andnot away from them. While it is possible to make the seal diameter largeenough so that the net force is negative, it is desirable to select aseal diameter which will provide a definite but mild force clamping thenozzle blades. While the magnitude of this force should be sufficient toassure no leakage between the actuator ring and the nozzle blades, itmust also be small enough so that the actuator ring will be able to moveaway from the nozzle blades to let particles pass.

In the case of high-pressure application, the difference in thepressure-break line on the blade side and the seal diameter on the outerface of the actuator ring becornes critical, inasmuch as this differencedetermines the clamping force on the blades. It has been found that in astructure in which the blades and actuator ring engage each otherthroughout their opposed surfaces, the position of the nozzle bladesnormally alters the diameter of the pressure-break lines on the bladeside.

Therefore, when a seal diameter is selected which will prevent anegative force with the nozzle blades in the closed position, thepressure-break line diameter on the blade side with the blades in openposition will shift, exposing a greater area to high pressure, therebycreating an excessive clamping force. The high clamping force with theblades in the open position makes it difficult if not impossible, toshift the actuator ring and causes the unlubricated sliding surfaces togall. Accordingly, it is a purpose of the present invention to controlthe pressure-break line diameter on the blade side of the actuator ringso that the diameter of the pressure-break line will remain fairlyconstant regardless of the position of the blades, whereby it ispossible to locate the seal diameter so that a small but definite andpractically constant clamping force will prevail between the rings andthe nozzle blades.

It is another object to provide an improved turbine having a variablenozzle assembly formed of spaced rings having a plurality of pivotednozzle blades therebetween wherein the actuator ring is free to moveaxially along a low-friction sealing surface whereby a slight clampingpressure may prevail between the rings and the nozzle blades.

It is still another object to provide a turbine with an improvedvariable nozzle assembly having an axially movable actuator ring whichexerts a small but definite and constant clamping force between therings and the nozzle blades and in which a substantial portion of thearea of the blade side of the actuator ring shielded from pressure bythe blades is relieved so that regardless of the position of the bladesthe exposed area will remain virtually constant.

It is a further object to provide an improved variable nozzle turbinehaving improved nozzle blades to improve flow characteristics.

It is still a further object to provide an improved variable nozzleturbine having an actuator ring having a slot whereby tangential thrustgenerates an outward force to reduce friction and wherein there is astop to prevent over-expansion.

SUMMARY OF THE INVENTION The turbine assembly of the present inventionis comprised of a housing having a fluid inlet and a fluid discharge. Aturbine wheel is rotatabl mounted on the axis of said housing. Avariable nozzle assembly formed of a pair of coaxially-spaced,coaxially-mounted rings and a plurality of pivoted blades sandwichedtherebetween is positioned between the inlet and the turbine wheel. Thespacings between complementary portions of adjacent nozzle blades formthe throats for the nozzle passages. One ring is fixedly-mounted and theother ring, which is the actuator ring, is mounted for angular rotativemovement relative to the first. The actuator ring is sealinglymounted ona bearing and is capable of axial movement. A plurality offixedly-positioned pins extend outward from the fixed ring and theindividual blades are pivotally-mounted on said pins. An elongated slotis located in each blade and a pin extending from the actuator ring isengaged therewith, whereby relative movement of the actuator ring variesthe position of the blades. Means are provided to apply a tangentialthrust on the actuator ring to rotate it concentrically with the axis ofthe turbine and thereby vary the opening of the nozzles.

High pressure acts on the outer face of the actuator ring upstream ofthe seal and low pressure acts on the outer face downstream of the seal.The blade side of the actuator ring is also subject to opposingpressures. The pressure-break line on the blade side in prior structuresnormally changes with the relative position of the nozzles. Accordingly,means are provided by this invention to maintain said pressure-breakline practically constant regardless of the position of the nozzleblades so that the seal diameter may be so relatively located that aslight but definite pressure is exerted by the actuator ring on thenozzle blades thereby sealing the nozzle blades and directing all of thefluid through the nozzle passages while at the same time permitting easyadjustment of the actuator ring and permitting axial movement of theactuator ring away from the nozzles to prevent jamming.

One means of maintaining the diameter of the pressurebreak line constantis to taper the surface of the actuator ring outwardly from a lineadjacent but upstream of the pivot pins of the fixed ring so that thereis a space between the blades and the opposing face of the actuatorring, whereby approximately the same area is subject to high pressureregardless of the position of the blades. If desired, the opposing wallof the fixed ring will also he correspondingly tapered. Althoughmovement of the tip portions of the blades does not make as large achange in area and the pressure in such region is low pressure, furtherequalization may be accomplished by discontinuous tapering of thedownstream region of the blade face of the actuator ring.

The actuator ring may he a solid ring mounted on lowfriction bearing andsealing surface or, if desired, the actuator ring may be a slotted ringhaving a stop to prevent over-expansion. In the latter case, atangential thrust is applied to the one side of the slot to providerotation and thereby alter the spacing of the blades, such thrustgenerates an outward radial thrust which reduces friction. Also, thefaces of the blades opposing the turbine wheel may be at an acute angleto the axis of the wheel to reduce friction of the flow through thenozzle passages.

BRIEF DESCRIE'TION OF THE DRAWINGS FIG. 1 is a cross section through avariable nozzle turbine constructed in accordance with the presentinvention, the section being take generally along the axis of theturbine;

FIG. 2 is an enlarged cross-sectional view to show the details ofconstruction through a variable nozzle blade;

FIG. 3 is a plan view of a portion of the variable nozzle assemblyshowing the nozzle blades in the open position in full line and inclosed position in dashed line;

FIG. 4 is a diagrammatic view of the nozzle blades illustrating thedifference in pressure-break line between the open and closed positionin prior art construction;

FIG. 5 is a view similar to FIG. 4 illustrating the difference inpressure-break line when the present invention is utilized;

FIG. 6 is a view similar to FIGS. 4 and 5 illustrating furtherimprovements in retaining the pressure-break line position constant;

FIG. 7 is a cross section view having an alternate nozzle bladeconstruction;

FIG. 8 is an isometric view illustrating the nozzle blade disclosed inH6. 7; and

FIGS. 9 and 10 are plan views illustrating alternate constructions forthe cam ring.

DESCRIPTION OF THE PREFERRED EMBODIMENT Referring now to the drawings,and in particular FIG. 1, there is illustrated a variable nozzle turbinehaving a housing 12 provided with a fluid inlet 14 and an axial fluiddischarge 16. Between the inlet and discharge is a turbine wheelcompartment 18 in which is located a turbine wheel 20 mounted on a shaft22 which extends through a sealed opening 24 in a closure member 26which sealingly closes an opening in housing 12. The turbine wheel 20 isprovided with a plurality of radially- .axially extending passages 28which are designed to receive fluid from inlet 14 and direct it throughturbine wheel 20 for discharge through axial discharge 16. A variablenozzle assembly 29 which controls the entry of fluid from the inletsurrounds turbine wheel 20.

The variable nozzle assembly 29 is formed of a pair of rings 30 and 32and a plurality of individuallypivotable nozzle blades 34 which aresandwiched between the two rings. The nozzle blades 34 are so mountedthat the spacings between complementary portions of adjacent bladesdefine throats forming nozzle passages 33. The ring 30 is fixedlyattached to the housing. The actuator ring 32 is axially-spaced fromfixed ring 30 and is mounted concentrically about a cover plate 36attached to the housing. Suitable means for providing rotative movementto actuator ring 32, such as an actuator rod 35, is provided. The coverplate forms a portion of axial discharge 16. A rotating seal 37 isprovided between the outlet end of passages 28 and a surroundingcircumferential portion 38 of cover plate 36. Another seal 39 is formedon the inner side of turbine wheel 26 between the wheel and acylindrical portion 41 of closure 26. Accordingly, all fluid whichenters through variable nozzles 33 is directed through turbine wheelpassages 28 and exits through the axial discharge 16.

Actuator ring 32 is mounted on a cylindrical bearing ring 40 of coverplate 36. The surface of the bearing ring opposing actuator ring 32 isprovided with a groove 42 in which is positioned a wafer-type springexpansion ring 44 and a seal ring 46 which may be formed of alow-friction material such as polytetrafluorethylene, com- .monlyreferred to by its trademark name Teflon or other suitable material.Seal ring 46 bears against inner circumferential surface 48 of actuatorring 32. If desired, inner surface 48 may be of stepped design toprovide a narrow metal-to-metal bearin contact between the actuator ringand the bearing surface.

As can be seen in FIG. 2, each blade 34 is pivoted on a pivot pin 50journaled in fixed ring 30. Each individual blade is provided with anelongated slot 52 in which a cam pin 54 extending from actuator ring 32is engaged. Accordingly, any rotative movement of actuator ring 32results in a movement of cam pin 54 which, through slots 52, results ina translating movement, varying the position of blades and changing therelative position of cam end 53 of one blade relative to tip end 55 ofan adjacent blade whereby the size of the nozzle passage 33 between twoadjacent blades is varied.

As previously mentioned, the space between actuator ring 32 and bearingring 40 is sealed by seal ring 46 and, as can be seen, actuator ring 32is free to move axially along cylindrical surface 40 of cover plate 36.There is a pressure drop through nozzles 33 and high pressure actsagainst outer face 56 of actuator ring 32 upstream of seal diameter SDand low pressure acts against that portion of face 56 which isdownstream of seal diameter SD. An opposing force prevails on nozzleblade face 58 of actuator ring 32. The force of blade side 58 is usuallysomewhat less than the force on outer face 56 whereby actuator ring 32will be urged gently against blades 34 and not away from them. Whileseal diameter SD may be made so large that the net force clamping thenozzle blades is negative, it is desirable to select a seal diameterwhich will provide a definite but mild force clamping the nozzle blades.The magnitude of this force must be small so that actuator ring 32 willbe able to move away from nozzle blades 34, if necessary, so as to letsome particles pass, if any occur in the gas being processed, to keepthem from becoming caught between the ring and blades and causing wearor malfunction.

In high-pressure applications, the difference in pressurebreak line onthe blade side and seal diameter SD on the opposite side of the actuatorring becomes critical since this difference determines the clampingforce on the blades.

It has been found that the position of nozzle blades 34 alters theaverage pressure-break line on the blade side. This is due to the factthat when there is a positive clamping force the blades will shield someof the area of the blade side of the actuator ring exposed to pressureand the amount shielded in the one position will differ from the amountshielded in the other positions. Thus, the average area of the bladeside subject to high pressure when the blades are in the open positionis much smaller than when the blades are in the closed position.Therefore, if a seal diameter is selected to prevent a negative force inthe closed position an excessively high clamping force will be createdwhen the blades are in the open position. T hehigh clamping force makesit diflicult, if not impossible, to shift the actuator ring and maycause the unlubricated sliding surfaces to gall.

Referring now to FIG. 4, several typical nozzle blades are shown pivotedon pivot pins 50. They are shown in solid outline in the open positionand in dashed outline in the closed position. As can be seen, the nozzleblades are moved from one position to the other by the circumferentialtranslating movement of cam pin 54 in slots 52. The solid line PBD showsthe estimated average pressure locus or pressure-break line between thehigh pressure and the low pressure areas on the blade side of theactuator ring for the open position, and dashed line PBC shows theaverage pressure locus or pressure-break line for the closed position.If the area enclosed by these two lines, as shown in FIG. 4, differs byabout /2 sq. in. per blade and if the pressure difference is 1000p.s.i., the blade in the open position is clamped with a force of 500lbs. greater than in the closed position. If the coefiicient of frictionis of the order of .8 and acts on both sides of the blade, it will takea force of the order of at least 800 lbs. to move one blade. For 16blades, this is an unacceptable load for a control device and metalsthis heavily loaded usually will not stand a dry rub between themwithout damage.

The reason for the difference in the pressure-break lines between theopen and the closed position is that as cam end 53 of the blade risesinto the upper high pressure zone it brings the average pressure-breakline with it, whereas on tip end 55 of the blade the change, although inthe opposite direction, is very small so that the net change issubstantially outwardly.

In order to minimize the difference in the clamping load between theopen and the closed position, the present invention brings the twopressure-break lines closer together. This is accomplished by relievingcertain portions of the mating faces of actuator ring 32 and blades 34.As can be seen in FIG. 2, surface'60 of actuator ring 32 and surface 62of fixed ring 30 are tapered outward from a circle adjacent the upstreamside of pivot pins 50. Therefore, the blade side of actuator ring 32 issubject to the high pressure outside of the line PB at all timesindependent of the position of the nozzle blades. At the same time, aseal is still established bet-ween the two rings and the blades in thearea'around the pivot pins which surfaces are not relieved. Accordingly,as can be seen in FIG. 5, the average diameter of the pressurebreak linePBD for the open position and the pressurebreak line PBC for the closedposition are almost the same. The cam end 53 of end blade 34 still risesinto the high pressure area as the nozzles open, but does not bring thepressure break-line with :it because the taper on the actuator ringexposes it to the high pressure down to a line almost through, but justoutside of, pivot pins 50. As can be seen in FIG. 5, the area betweenthe lines P'BD' and PBC' is much less than that between PBD and PEG inFIG. 4 and this degree of balance is usually sufficiently close toprevent galling and maintain the force required to rotate the actuatorring at a satisfactory level. Although it is not necessary to have acorresponding taper on fixed ring 30, it is desirable to do so to assurebalanced forces on both sides of the nozzle blades. Such tapering makesthe system symmetrical and eliminates any tendency for the nozzle bladeto be pressed against the fixed ring by some unusual pressuredistribution.

If complete balance is desired, it can be accomplished by furtherrelief, such as by tapering surface 64 of actuator ring 32 and surface66 of fixed ring 30 downstream of pivot pins 50. This additional reliefas shown in FIGS. 2 and 6 is discontinuous covering only the areadefined by X, Y and Z, that is to say, the taper surface isdiscontinuous between lines X and Z. If the actuator ring had a fullytapered surface, some of the effectiveness would be lost since therewould be blow-by under the tip of the nozzle blade which would beundesirable. As can be seen in FIG. 6, the estimated pressure-break lineP"B"C" for the closed position is less than the pressurebreak line P"BD"for the open position. This is the reverse of that depicted in FIG. 4and demonstrates the possibility of reaching a point of equality betweenthe two pressure-break lines.

Accordingly, it can be seen that by relieving the influence of the camends of the nozzle blades when they are moved from one position toanother the pressurebreak line remains fairly constant; therefore sealdiameter SD on the outer side of actuator ring 32 can be accuratelydetermined to provide the desired clamping force. While tapered surfacesare shown as preferred means for relieving the pressure caused byshifting of the nozzle blades from one position to another, other meanssuch as a circumferential groove may also be utilized. Any means capableof equalizing the shielding of the blade side of the actuator ring frompressure regardless of position of the nozzle blades may be utilized.

Although the tapered surfaces provide a passage which permits theupstream high pressure to be in constant communication with the bladeside of the actuator ring and the opposing faces of the blades, suchtapered surfaces terminate upstream of the circle formed by the fixedpivot pins. Therefore, provided a positive clamping force exists, theactuator ring will be forced against the blades and the blades againstthe fixed ring and a seal will be established in the area adjacent thepivot pins and there will be no blow-by.

The nozzle blade illustrated in FIGS. 1 through 5 has its end portion 55parallel to the axis of turbine wheel 20, Le, the surface of the bladeclosest to the axis of the turbine wheel is generated by a line whichmoves parallel to the axis of the turbine wheel. In nozzle blade 80,illustrated in FIGS. 7 and 8, there is an end portion which is at anacute angle to the axis of turbine wheel 86 and which, if projected,will intersect the axis of the turbine wheel. The surface 84 of blade80, which surface corresponds to the underneath surface of the forms ofFIGS. 1-6, is generated by a line which is positioned at an angle to theaxis of the turbine wheel. The angle at the end 88 is an acute angle.The underneath surface may be generated by such an angularly-positionedline whose angularity from end 88 to the heel of the blade is eitherconstant or decreases in angularity as it moves from end 88 to the heelof blade 80.

In the form shown in FIG. 8, the surface is generated by a generatingline of changing angularity as described above. The angularly-positionedsurface 84 of the blade has an important advantage. The fluiddischarging from the nozzle into passageways 90 of wheel 86 is notsubjected to any material friction loss since it is not subjected toflow over frictional surfaces between the surface 84 and the passagewayentrances. The shroud 92 and the disc 94 of turbine wheel 86 actingtogether with the vanes from passageways 90 through turbine wheel 86whose inlet at the nozzle has a direction with a substantial axialcomponent.

The construction described above minimizes the area of surface exposedto the high-velocity stream from the nozzle to the entrance of turbinewheel 86.

The actuator ring 32 illustrated in FIGS. 1 through 6 may be a solidring riding on the low-friction Teflon seal ring 46. Inasmuch as theseal ring is formed of lowfriction material, the frictional drag duringrotative movement of the ring is kept to a minimum. However, if desired,the actuator ring may be provided with a slot and the tangential thrustmade at one side of the slot to rotate the actuator ring. The tangentialthrust generates an outward radial thrust on the actuator ring andcauses the friction between the actuator ring and the bearing ring to bereduced resulting in easy rotative movement of the actuator ring. Theslot may be merely a slit in the ring, as for example a saw cut, or itmay take the forms shown in FIGS. 9 and 10, in which case there is astop to prevent overstrain of the ring when the tangential pressure isapplied. This prevents any overstrain which could possibly cause thering to fail.

As shown in FIG. 9, the slot 100 is formed to produce an internal stop.The slot is formed with a bore 101 to receive an actuator member 102.The slot 100 cuts the ring across the bore and extends circumferentiallyat 104, makes a reverse U-bend at 105 and extends radially across ring106; however, being stepped at 107 and 108 to form an offset labyrinthout which minimizes any fluid leakage through the slot. The reverseU-bend at 105 acts as an internal stop to limit expansion of ring 106.

In the form shown in FIG. 10, a slot 110 extends radially across ring111, being stepped at 112 and 113 to form the labyrinth cut whichminimizes fluid leakage. The slot has a bore 114 in which an actuatormember 115 is located. The ring is provided with an externallypositioneddog 116 and stop 117 positioned on each side of a slot 110 forming aneffective stop against undue expansion of the ring. Therefore, as can beseen, the slots are so designed that the actuator ring cannot be undulyexpanded, therefore preventing any overstrain of the actuator ring.

From the foregoing it will be seen that this invention is one welladapted to attain all of the ends and objects hereinabove set forth,together with other advantages which are obvious and which are inherentto the apparatus.

It will be understood that certain features and sub-' combinations areof utility and may be employed without reference to other features andsubcombinations. This is contemplated by and is within the scope of theclaims.

As many possible embodiments may be made of the invention withoutdeparting from the scope thereof, it is to be understood that all metterherein set for.h or shown in the accompanying drawings is to beinterpreted as illustrative and not in a limiting sense.

I claim:

1. A turbine assembly comprising a housing, a fluid inlet and a fluidoutlet in said housing, a turbine wheel rotatably mounted on an axis insaid housing, a pair of rings coaxially mounted about said axis in saidhousing, the first of said rings being fixedly-mounted in said housingand the second of said rings mounted on a bearing for angular rotativemovement relative to the first of said rings and movable axially alongthe bearing, a seal between said second ring, which is an actuator ring,and its bearing, a plurality of adjacent nozzle blades mounted betweensaid rings at spaced intervals about said rings, the spacing betweensaid complementary portions of adjacent nozzle blades forming a throatfor a nozzle passage, means including said actuator ring for alteringthe spacing between said complementary portions of said nozzle blades,there being an area exposed to high pressure on the outer side of saidactuator ring upstream of said seal, and an area exposed to low pressuredownstream of said seal and an area exposed to upstream high pressureand an area exposed to downstream low pressure on the blade side of saidactuator ring, a potion of the blades moving into the high-pressure areain their nozzle open position, and said blades and actuator being formedto provide spaces therebetween and expose to high pressure the areas ofsaid actuator ring in the high pressure area into which said blades somove for maintaining the area on the blade side of the actuator ringsubject to high pressure constant within predetermined limits regardlessof the position of the blades.

2. The turbine assembly specified in claim 1, wherein there is a passagebetween the blade side of the actuator ring and the opposing face of theblades which is subject to upstream high pressure, such passagemaintaining the area of the blade side of the actuator ring subject tohigh pressure constant within predetermined limits regardless of theposition of the blades.

3. The turbine assembly specified in claim 1, wherein the blades arepivotally mounted on pivot pins on the fixed ring, the surface of theblade side of the actuator ring is tapered outward of a seal areacircumscribed by a circle adjacent the upstream side of the pivot pinswhereby pivotal movement of the blades will not materially vary theeffective area subject to high pressure.

4. The turbine assembly specified in claim 3, wherein the surface of thefixed ring corresponding to the tapered surface of the actuator ring iscorrespondingly tapered.

5. The turbine assembly specified in claim 3, wherein the surface of theblade side of the actuator ring downstream of the seal line is providedwith outwardly tapered portions so that movement of the blades in o andout of such portions will not materially vary the effective area subjectto low pressure.

6. The turbine assembly specified in claim 5, wherein the surfaces ofthe blade side of the fixed ring upstream and downstream of the pivotpins are provided with tapered portions corresponding to the taperedportions on the actuator ring.

7. The turbine assembly specified in claim 1, wherein the actuator ringhas a radially-extending slot and the blade space altering means appliesa tangeniial thrust at one side of said slot.

8. The turbine assembly specified in claim 7, wherein said radial slotis in the form of a labyrinth passage to minimize flow leakagetherethrough.

9. The turbine assembly specified in claim 8, wherein said actuator ringhas stop means cooperating with the slot to prevent undue expansion ofsaid actuator ring by the application of the tangential thrust.

10. The turbine assembly specified in claim 9, wherein said stop meansis formed of an externally-positioned first stop member on the outerperiphery of the actuator ring on one side of the slot and a secondmember on the outer periphery of the actuator ring on the other side ofthe slot having a recess cooperating with the first stop member.

11. The turbine assembly specified in claim 9, wherein said stop meansis formed by the slot extending radially inward from the outer edge,then circumferentially, and then having an inverted U-bend, one leg ofwhich extends toward the inner edge.

12. The turbine assembly specified in claim 1 wherein a plurality ofpivot pins extend from the fixed ring and the individual nozzle bladesare pivoted on said pins, the individual nozzle blades and the actuatorring having inter-engaging, elongated slots and cam pins which engagewith said slots respectively whereby rotation of the actuator ringcauses the nozzle blades to pivot about their pivot pins therebychanging the spacing between complementary portions adjacent nozzleblades, the surface of the nozzle blades adjacent to the turbine wheelbeing parallel to the axis of the turbine wheel.

13. The turbine assembly specified in claim 1, wherein a plurality ofpivot pins extends from the fixed ring and the individual nozzle bladesare pivoted on said pins, the individual nozzle blades and the actuatorring having inter-engaging, elongated slots and cam pins which engagewith said slots respectively whereby rotation of the actuator ringcauses the nozzle blades to pivot about their pivot pins therebychanging the spacing between complementary portions of adjacent nozzleblades, the surface of the nozzle blade adjacent to the turbine wheelbeing at an acute angle to the axis of the turbine wheel.

14. A turbine assembly comprising a housing having a high pressure fluidinlet, and a low-pressure fluid discharge, a turbine wheel rotatablymounted on an axis in said housing, a pair of rings coaxially mountedabout said axis, one of said rings being fixedly mounted in said housingand the other of said rings being an actuator ring mounted for rotativemovement relative to the fixed ring, a plurality of adjacent nozzleblades mounted on and between said rings and movable to provide largeror smaller nozzle openings, said actuator ring having an outercircumferential surface in fluid communication with high pressure fromsaid inlet and an opposite inner ring surface, a cylindrically outwardlyfacing bearing surface in said housing, the inner surface of saidactuator ring opposing said bearing surface, an annular recess in saidbearing surface, a low-friction sealing ring in said recess sealinglyengaging said inner ring surface of said actuator ring, the outer facesurface of said actuator ring being in fluid communication with saidinlet upstream of said seal and the outer surface of said actuator ringdownstream of said seal being in fluid communication with the pressuredownstream of the nozzle blades, the blade side of said actuator ringalso having an area exposed to said pressure from said inlet wherebythere are opposing axial pressures on said actuator ring, and the bladeside of said actuator ring and the opposing face of the blades beingexposed to said high pressure whereby a change in position of the nozzleblades will not materially affect the area on the blade side of theactuator ring subject to the high pressure.

15. The turbine assembly specified in claim 14 wherein the blades aremounted on pivot pins and the upstream surface of the blade side of theactuator ring is tapered outward of a seal line extending through thepivot pins to vent said area so that movement of the blades will notmaterially vary the effective area subject to high pressure.

16. The turbine assembly specified in claim 15, wherein the upstreamsurface of the fixed ring opposing the tapered surface of the actuatorring is correspondingly tapered.

17. The turbine assembly specified in claim 14 wherein the downstreamsurface of the blade side of the actuator ring is provided with taperedportions so that movement of the blade into and out of such portionswill not materially vary the effective area subject to low pressure.

18. The turbine assembly specified in claim 17, wherein the upstream anddownstream surfaces of the blade side of the fixed ring opposing thetapered surfaces of the actuator ring are provided with correspondinglytapered portions.

References Cited UNITED STATES PATENTS 1,197,761 9/1916 Pfau 253-1223,232,581 7/1963 Swearingen 253122 3,243,159 3/1966 Hefler et al 253-122EDWARD L. MICHAEL, Primary Examiner

